PIV Measurements in the Impeller and the Vaneless Diffuser of a Radial Flow Pump in Design and Off-Design Operating Conditions

Transcription

1 G. Wuibaut G. Bois ENSAM, 8, Bd Louis IV, Lille, Cedex, France P. Dupont Ecole Centrale de Lille, BP 48, Villeneuve d Ascq Cedex, France G. Caignaert ENSAM, 8, Bd Louis IV, Lille, Cedex, France Guy.Caignaert@lille.ensam, fr M. Stanislas Ecole Centrale de Lille, BP 48, Villeneuve d Ascq Cedex, France PIV Measurements in the Impeller and the Vaneless Diffuser of a Radial Flow Pump in Design and Off-Design Operating Conditions This paper presents and discusses the results of an experimental program that has been made on an air test rig of a radial flow pump. The tested impeller is the so-called SHF impeller. Many experimental data have already been produced (tests in air and in water) on that geometry and these results are still used as databases for the validation of CFD codes. For the present study, an air test rig has been chosen for optical access facilities and measurements were realized with a vaneless diffuser. The 2D Particle Image Velocimetry technique has been used and measurements of flow velocities have been made simultaneously in the outer part of the impeller and in the vaneless diffuser. Measurements have been realized in five planes, in the hub to shroud direction, for various relative flow rates (design and off-design operating conditions). First, the paper focus on the evolutions of the phase averaged velocity charts in the impeller and the diffuser. Limitations of the phase averaging technique clearly appear in the very low partial flow rates and this will be related to previous pressure measurements analysis establishing the occurrence of rotating stall within the impeller for such operating conditions. The paper also proposes an analysis of the rates of fluctuations of the velocity charts and the evolutions in the various measuring planes as the relative flow rate becomes lower. DOI: / Introduction The design of pumps is mainly based on steady flow assumptions in runner, vaneless, and vaned diffusers. This kind of approach is suitable for design operating conditions of classical turbomachinery geometries. However, to understand and take into account wide operating ranges in pump designs, it is necessary to improve the knowledge of unsteady effects and rotor-stator interactions. Experimental data are more and more required in order to calibrate new design techniques including the unsteady character of the flow. For a better optimization of the pump design, numerical simulations of internal flow are now proposed, including a coupling between the different parts of the machines, and it is necessary to validate these methods to define the validity range of the various kinds of approach. The present paper refers to a first series of results of an experimental program that has been realized on an air test rig of a radial flow pump optical access facilities. These first results are related to a pump configuration with a vaneless diffuser and no volute. The tested impeller is the so-called SHF impeller for which many experimental and numerical investigations have already been made and described in previous works references 1 8. Up to now, these results mainly refer to static and dynamic pressure measurements and to hot wire s anemometers or Laser Doppler Velocimetry. Particle Image Velocimetry appears to be very useful for a better understanding of phenomena associated to rotor-stator interactions In the present paper, PIV has been used for the analysis of flow Contributed by the Fluids Engineering Division for publication in the JOURNAL OF FLUIDS ENGINEERING. Manuscript received by the Fluids Engineering Division July 27, 2001; revised manuscript received April 4, Associate Editor: J. Katz. velocities in the outer part of the impeller and the vaneless diffuser. Results in design and off-design operating conditions are presented and analyzed. 1 Experimental Setup The air test rig has been already described in references 13,14, as well as the PIV measurement device and data acquisition procedure. The impeller main characteristics are the following: outlet radius: R mm tip inlet diameter: mm outlet width: 38.5 mm outlet blade angle: 22.5 deg measurement from the peripheral direction mean blade thickness: 9 mm number of blades: Z 7 speed of rotation: up to 2500 rpm design flow rate: QN m 3 /s at 1710 rpm The outlet part of that impeller is characterized by a 2D design. The vaneless diffuser has been designed with an inlet diameter equal to 515 mm, an outlet diameter equal to 575 mm and a constant width B3 equal to 39 mm. As it can be seen in Fig. 1, PIV measurements have been made in planes perpendicular to the pump axis of rotation. Each plane is defined by its axial position B with a reference to the hub part of the diffuser. Through the transparent shroud parts of the impeller and the diffuser, the PIV camera observes the Poly Ethylene Glycol particles which are injected into the flow with a smoke generator. The PIV System is based on cross correlation technique. An electronic box has been used in order to synchronize the CCD camera and YAG pulsed laser with the impeller rotation. One instantaneous velocity field is measured Journal of Fluids Engineering Copyright 2002 by ASME SEPTEMBER 2002, Vol. 124 Õ 791

2 Fig. 1 each two complete revolution of the impeller. The tests have been made with the following conditions: energy of each laser pulse: 250 mj pulse duration: 10 ns delay between two pulses: 35 s The camera is a Kodak camera with pixels. The experimental database contains results for the following conditions: speed of rotation: 1710 rpm relative flow rates Q/QN: 0.26, 0.45, 0.63, 0.91, 1.02, 1.61 five measurements planes in the hub to shroud direction: B/B3 0.12, 0.25, 0.50, 0.74, 0.87 two different views: view 1 covers the outlet throat of the impeller and covers one blade trailing edge. The two views are symbolically presented in Fig. 2. For each operating and measuring conditions, 230 sets of two successive images have been registered. Each set of two successive images is then analyzed in order to get one velocity map, using the image cross correlation technique. In order to improve the quality of the results, a background view has been subtracted on each PIV image. All sets of two consecutive images have then been analyzed with or pixels elementary cells. For the analysis of operating conditions near design conditions, an advanced digital interrogation technique has been used by using a window offset 15. Finally, a mask has been used for each view in order to eliminate vectors in non-fluid zones. Such a procedure allows for the determination of two components of absolute fluid velocities in the measuring plane, defined in the frame of each view. A post-treatment procedure has been developed in order to locate the point of the pump axis of rotation in the view frame so that it becomes possible to define the radial and peripheral directions in each point of a view and then obtain the absolute or Fig. 2 Definition of a measurement plane Definition of the two views relative velocities and their radial and peripheral components or the flow angles between the absolute velocity and the peripheral direction and between the relative velocity and the peripheral direction. In every measurement plane and for each view, 230 instantaneous velocity maps are available for every operating condition. Measurements have been made for the same position of the impeller and synchronous averages have been calculated for each series of 230 instantaneous velocity fields. According to the various sources of uncertainties during the measurement procedure, the relative uncertainty of the instantaneous flow velocity, in each point of the measurement grid, has been estimated to 2%. In each point of the measurement grid of a view, the instantaneous components of the absolute flow velocity are defined as u(x,y,t) and v(x,y,t). A good estimation of the synchronous averaged velocity in that point can so be obtained with an arithmetical average over the 230 data Eq. 1. ū x,y u x,y,t 0 i t (1) v x,y v x,y,t 0 i t Using the Reynolds decomposition model Eq. 2 it becomes possible to obtain a kinetic turbulence energy K from the calculation of the arithmetical average of the square of velocity fluctuations u (x,y,t) and v (x,y,t) Eq. 5. u x,y,t ū x,y u x,y,t v x,y,t v x,y v x,y,t (2) u 2 x,y u 2 x,y,t 0 i t (3) v 2 x,y v 2 x,y,t 0 i t (4) K x,y 1 2 u 2 x,y v 2 x,y (5) Tu x,y K x,y R 2. 2 Phase Averaged Velocity Components Blade to blade evolutions of C u /U 2, nondimensional mean absolute velocity peripheral components, are shown for a particular radius inside the impeller (R/R ) in Fig. 3. That particular section has been chosen for further comparisons with LDV measurements and calculations from 15 and 7. The various parts of Fig. 3 correspond to four measuring planes between hub and shroud. In each figure, the evolutions have been plotted for various operating conditions, from Q/QN 0.45 to Q/QN Each figure contains averaged results issued from the two views with some amount of overlapping. That overlapping in the posttreatment shows a very good coherence between the results in the two views, for the flow rates Q/Qn 1.61 to In Fig. 3, for the part load conditions Q/Qn 0, 45 and 0.63, close to the position angle / 2 /7, the gap in the curves can be explained by the fact that the coherence between two view s data is decreasing due to the unsteady character of the flow. In Fig. 3, for the position angle between 0.1 and 0.25, the gap in curves is due to a lack of informations between the two views as shown in Fig. 2. First of all, it is interesting to make a comparison between integrated values of R. Cu issued from the PIV measurements with the ones deduced from the available overall performances of the impeller in terms of internal head 1. No PIV measurements can be obtained at impeller outlet (R R2) because of flare due to (6) 792 Õ Vol. 124, SEPTEMBER 2002 Transactions of the ASME

3 Fig. 4 Overall mean values of the peripheral component of flow velocities at impeller outlet Fig. 3 Evolutions of CuÕU 2 in the blade to blade direction for various axial positions and a fixed radius the gap between the rotating and the stationary parts of the pump. In order to obtain the mean value of R. Cu at the impeller outlet, both results inside the impeller and in the vaneless diffuser have been extrapolated to the outlet impeller radius. The integrated local values of RCu from PIV measurements are based on only five hub to shroud positions. So, it is evident that the obtained value of RCu may be wrong. It has been decided to plot the ratio Cu /Cun with PIV, making the assumption that the relative error could be of the same order of magnitude for different mass flow conditions. For design conditions, the total head coefficient obtained by the PIV measurements and the one deduced from torque measurements are quite in agreement within 0.3%. For the other operating conditions, it can be seen in Fig. 4 that the PIV results closely correspond to the overall curve of reference 1 except for a nondimensional flow rate equal to Looking at the blade-to-blade evolution of C u /u 2, it can be seen that they are quite similar, for each flow rate, whatever the axial position B/B3 between and In the shroud region, levels and gradients are modified, especially for operating conditions corresponding to a nondimensional flow rate equal to Similar considerations can be extracted from the observation of the radial velocity components. A clear modification can be observed for B/B3 0.74, below the design flow rate in the suction side of the impeller blade. It appears that the impeller suction side shroud corner is dominated by an important relative velocity gradient and an accumulation of low momentum fluid. For the rest of the blade passage section, the relative velocity is mainly uniform along the blade height while it regularly increases from the pressure side to the suction side as can be seen in Fig. 5 for design conditions, corresponding to a classical nonviscous flow pattern. These kinds of flow structures have been already well reported by Eckardt 17 and Ubaldi 18 for machines comparable to the present one. For the present case and for design flow operating conditions, this particular flow structure seems to start below values of R/R inside the impeller. This location is situated after the axial to radial bend of the impeller where important decrease of turbulence due to Coriolis effects occur in the suction side shroud corner as explained by Eckardt. The so-called wake can be easily detected in the Fig. 5 by the comparison between the isovalues curves of the relative mean velocity in for axial positions B/B and B/B , starting from the shroud compared with the one obtained at mid-height (B/B3 0.5). When the flow goes toward the blade outlet section, suction blade curvature continues to act on the fluid with higher values of Coriolis effects due to rotation. Wake is so carried toward the mid-height on the suction side. The wake coming from the pressure side can also be seen in the vaneless diffuser channel where its location tends to be transported toward the mid blade to blade passage. Journal of Fluids Engineering SEPTEMBER 2002, Vol. 124 Õ 793

4 Fig. 6 Rates of velocity fluctuations BÕB3Ä0.5 QÕQnÄ1.61 on the losses over the entire blade suction side. It also has to be noted that, for Q/QN 0.63, the potential flow organization outside the wake region does not exist anymore. This aspect may be related to the modification already described in Section 2 concerning the tangential velocity gradient modifications. It must be remembered 1 that the impeller inlet recirculation critical flow rate has been experimentally determined for Q/QN So, for Q/QN 0.63, the inlet impeller recirculation may be more extended; this can also explain the velocity field modifications observed inside the impeller and the vaneless diffuser. The evolutions of nondimensional velocity fluctuation rate Tu(x,y) Eq. 6 are presented in Figs for different flow rates. These figures relate to measurements at mid-height. In design conditions Fig. 7, the velocity fluctuations are very low within the outer part of the impeller: they present less than 5% of the peripheral velocity based on the outlet impeller radius R 2. In this region, the fluctuations seem to be equivalent in the pressure and suction sides vicinity, with higher values in the suction side outlet region of the impeller. For Q/QN 1.61, fluctuations on the suction side are lower but the level reaches 6% on the pressure side. In the diffuser outlet, Fig. 5 Isovalues of non-dimensional mean relative velocities 3 Off-Design Conditions Always for the same radial position R/R and for partial flow operating conditions, the wake structure grows, reaches the blade mid-height for Q/QN 0.91 and fill the entire blade suction side for Q/QN For Q/QN 0.45, the wake structure is no longer detected since the effects of the positive incidence at the blade leading edge of the impeller blades lead to dominant effects Fig. 7 Rates of velocity fluctuations BÕB3Ä0.5 QÕQnÄ Õ Vol. 124, SEPTEMBER 2002 Transactions of the ASME

5 Fig. 8 Rates of velocity fluctuations BÕB3Ä0.5 QÕQnÄ0.91 Fig. 10 Rates of fluctuations BÕB3Ä0.5 QÕQnÄ0.45 the fluctuation rates become higher nearly 6% as the absolute flow velocities are becoming much smaller: this can also be attributed to the influence of the sudden expansion at the diffuser outlet and to flow stability limits, with higher fluctuations near the shroud as already described in 16 for a vaneless diffuser with a higher outlet radius. Figure 11 a presents the same type of results, at mid-height also, but for the view 1. As it can be seen on that figure, some problems with light reflection and diffusion by the trailing edge of the blade have been encountered during the tests, and it has been decided to mask a part of the view in order to protect the camera. In that figure, it appears first that the rate of velocity fluctuations within the diffuser is higher up to 7% along what can be called the wake of an impeller blade. As shown in Fig. 11 b the spread of that wake zone appears to be wider near the hub. Second, a large region with very low fluctuation rates can be seen within the impeller blade to blade section near the suction side. This is probably related to the evolution of the low momentum part that develops inside the blade passage near the shroud section side corner as already pointed out in the previous section. Fig. 9 Rates of velocity fluctuations BÕB3Ä0.5 QÕQnÄ0.63 Fig. 11 a rates of fluctuations BÕB3Ä0.5 QÕQnÄ1.02 view 1; b rates of fluctuations BÕB3Ä0.256 QÕQnÄ1.02 view 1 Journal of Fluids Engineering SEPTEMBER 2002, Vol. 124 Õ 795

6 Fig. 12 Rates of fluctuations BÕB3Ä0.5 QÕQnÄ0.63 view 1 Fig. 13 Rates of fluctuations BÕB3Ä0.5 QÕQnÄ0.26 view 1 For lower flow rates, Figs 12 and 13, the rates of fluctuations regularly increase both in the vaneless diffuser 0.08 to 0.14 for Q/QN 0.63 and in the pressure side zone inside the impeller up to 10% for Q/QN 0.26 where large instabilities with reverse flows have been observed in relationship with the behavior of the vaneless diffuser in partial flow conditions. In these cases, the instabilities progressively propagate from the outer part of the diffuser to the outer part of the impeller pressure side. However, it has to be kept in mind that it is rather difficult to define correctly a phase averaged velocity chart for relative flow rates lower than 0.45 with only 230 instantaneous velocity charts. On the contrary, the suction side zone is not affected by the vaneless diffuser instabilities since the fluctuation rates remain low about 3%, except for Q/QN 0.63 where the levels reach locally 8% close to the suction side. This corresponds to a zone which has been already described by several authors: a new increase of turbulence may occur near the suction side of the blade, in association with the boundary layer separation that creates the so-called jet and wake configuration. Conclusion and Perspectives The PIV technique has been successfully applied to the instantaneous characterization of velocity distributions in a radial flow pump impeller associated with a vaneless diffuser for design and off-design operating conditions. Near design conditions, flow within each blade to blade passage can be considered as steady in the relative frame according to the axisymmetry of the casing vaneless diffuser without volute. The so-called jet and wake structure can also be observed inside impeller blade passage as well as in the vaneless diffuser, associated with local velocity fluctuations and/or instabilities coming from the diffuser itself. The pressure side region is affected by diffuser instabilities whereas the suction side is more dependent on inlet conditions and impeller flow developments. At very low flow rates (Q/QN 0.26), large instabilities develop in the vaneless diffuser and even propagate in the outer part of the impeller with unsteady flow patterns in the relative frame. The available database still need the development of posttreatment procedures in order to become more useful especially for the validation of numerical approaches in off-design conditions. These developments and comparisons are still in progress. The technique is now used on the same test rig to study impellervaned diffuser interactions. Nomenclature R radius R 2 outlet radius of the impeller Z number of blades of the impeller N speed of rotation impeller angular velocity Q flow rate QN design flow rate B3 constant width of the vaneless diffuser B distance between the laser sheet and the hub (x,y) coordinates of a point in the frame grid t 0 time of the first sample t duration of two turns of the impeller u(x,y,t) absolute instantaneous velocity x component v(x,y,t) absolute instantaneous velocity y component ū(x,y) phase averaged velocity x component Eq. 1 v (x,y) phase averaged velocity y component Eq. 1 u (x,y,t) instantaneous velocity fluctuation x component Eq. 2 v (x,y,t) instantaneous velocity fluctuation y component Eq. 2 K(x,y) turbulent kinetic energy at location (x,y) Eq. 5 C u peripheral component of phase averaged absolute velocity for a constant B C u mass averaged peripheral velocity C un mass averaged peripheral velocity at nominal flow rate C r radial component of phase averaged absolute velocity U 2 U 2 R 2 W phase averaged relative velocity Fig. 5 angle References 1 Barrand, J. P., Caignaert, G., Canavelis, R., and Guiton, P., 1984, Experimental Determination of the Reverse Flow Onset in a Centrifugal Impeller, Proceedings of the First International Pump Symposium. Texas A&M University. 2 Barrand, J. P., Caignaert, G., Graeser, J. E., and Rieutord, E., 1985, Synthèse de Résultats an Air et en eau en vue de la Détection des Débits Critiques de Recirculation à l entrée etàla Sortie de la roue d une Pompe Centrifuge, La Houille Blanche No. 5, pp Caignaert, G., Desmet, B., Maroufi, S., and Barrand, J. P., 1985, Velocities and Pressure Measurements and Analysis at the Outlet of a Centrifugal Pump, ASME Paper 85-WA/FE-6. 4 Caignaert, G., Barrand, J. P, and Desmet, B., 1988, Recirculation at Impeller 796 Õ Vol. 124, SEPTEMBER 2002 Transactions of the ASME

7 Inlet and Outlet of a Centrifugal Pump, Proceedings of the Institution of Mechanical Engineers, ImechE , paper C337-88, pp Caignaert, G., and Morel, P., 1995, Mean Pressure Measurements Within a Centrifugal Pump Impeller at Partial Flow Rates, VDI Berichte, 1186, pp Caignaert, G., and Patricio, O., 1998, Pressure Fluctuations Within the Impeller and the Inlet Duct of a Centrifugal Pump, Hydraulic Machinery and Cavitation, World Scientific, ed, pp Combes, J. F., and Rieutord, E., 1992, Numerical and Experimental Analysis of the Flow in a Centrifugal Pump at Nominal and Partial Flow Rate, ASME paper 92 GT 248, 12 pages. 8 El Hajem, M., Morel, M., Spettel, F., and Bois, G., 1998, Etude de l Ecoulement Moyen en Sortie de roue d une Pompe Centrifuge roue SHF, La Houille Blanche, No. 7, pp Stanislas, M., Kompenhans, J., and Westerweel, J., 2000, Particle image velocimetry: towards industrial applications, Kluwer Academic Press, Amsterdam. 10 Sinha, M., and Katz, J., 2000, Quantitative Visualization of the Flow in a Centrifugal Pump With Diffuser Vanes. I: on Flow Structures and Turbulence, ASME J. Fluids Eng., 122, pp Sinha, M., Katz, J., and Meneveau, Ch., 2000, Quantitative Visualization of the Flow in a Centrifugal Pump With Diffuser Vanes. II: Addressing Passage- Averaged and Large-Eddy Simulation Modeling Issues in Turbomachinery Flows, ASME J. Fluids Eng., 122, pp Myers, K. J., Ward, R. W., and Bakker, A., 1997, A Digital Particle Image Velocimetry Investigation of Flow Field Instabilities of Axial-Flow Impellers, ASME J. Fluids Eng., 119, pp Wuibaut, G., Dupont, P., Caignaert, G., and Stanislas, M., 2000, Experimental Analysis of Velocities in the Outlet Part of a Radial Flow Pump Impeller and the Vaneless Diffuser Using Particle Image Velocimetry, XX IAHR Symposium, Charlotte, 6 9 Aug. 2000, paper GU Wuibaut, G., Dupont, P., Bois, G., Caignaert, G., and Stanislas, M., 2001, Analysis of Flow Velocities Within the Impeller and the Vaneless Diffuser of a Radial Flow Pump, 4th European Conference on Turbomachinery, Fluid Dynamics and Thermodynamics, March, 2001, Firenze, Italy, paper ATI-CST-047/ Raffel M., Willert C., and Kompenhans, J., 1998, Particle Image Velocimetry, Springer-Verlag, Berlin Heidelberg, p Johnston, J. P., and Eide, S. A., 1976, Turbulent Boundary Layers on Centrifugal Compressors Blades: Prediction of the Effects of Surface Curvature and Rotation, ASME J. Fluids Eng., 98, pp El Hajem, M., Akhras, A., Morel, R., and Champagne, J. Y., 2001, Rotor Stator Interactions in a Centrifugal Pump Equipped with a Vaned Diffuser, 4th European Conference on Turbomachinery, Fluid Dynamics and Thermodynamics, Mar. Firenze, Italy. 18 Ubaldi, M., Zunino, P., and Cattanei, A., 1993, A. Relative Flow and Turbulence Measurements Downstream of a Backward Centrifugal Impeller, ASME J. Turbomach., 115, pp Journal of Fluids Engineering SEPTEMBER 2002, Vol. 124 Õ 797